Dual clutch transmission with multiple range gearing

ABSTRACT

A vehicle powertrain comprising dual torque input clutches and a multiple speed transmission with ratio range gearing is disclosed. The range gearing provides increased gear ratio coverage. A transition between a low speed ratio range and a high speed ratio range is achieved with no torque interruption. A shift from one gear ratio to an adjacent gear ratio in a speed ratio range during a shift sequence is preceded by preselecting the adjacent gear ratio.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application is a continuation of U.S. application Ser. No.11/702,361 filed Feb. 5, 2007, now U.S. Pat. No. 7,621,839. Applicantclaims the benefit of that patent application.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The invention relates to geared, multiple-ratio power transmissionmechanisms having dual clutches to permit selective engagement anddisengagement of ratio change clutches for operation in a given ratiowhile preselecting ratio change clutches for an adjacent ratio during ashift sequence in separate speed ratio ranges.

2. Background Art

Conventional automotive vehicle powertrains typically havemultiple-ratio transmission mechanisms that establish power deliverypaths from an engine to vehicle traction wheels. Adequate powertrainperformance for heavy-duty or medium-duty powertrains requires more thanone range of gear ratios so that the speed and torque characteristic ofthe engine will provide optimum traction wheel power throughout theoperating speed range for the vehicle. Ratio range gearing, therefore,typically is used with a transmission mechanism in such high torquecapacity powertrains so that sufficient overall gear ratio coverage isprovided. The overall gear ratio coverage will be equal to the productof the number of gear ratios for the transmission and the number of gearratios in the ratio range gearing.

One class of transmission mechanisms in automotive vehicle powertrainsincludes countershaft gearing having power delivery gears journaled on atransmission mainshaft and countershaft gear elements journaled on acountershaft arranged in spaced, parallel disposition with respect tothe mainshaft. The countershaft gear elements typically mesh with thegears mounted on a mainshaft axis. Operator-controlled clutches, whichmay be either positive drive dog clutches or synchronizer clutches,selectively connect torque transmitting gears on the mainshaft axis tothe mainshaft, thereby establishing a power delivery path from theengine to the traction wheels. In other countershaft transmissionmechanisms, the dog clutches or the synchronizer clutches may be mountedon the countershaft axis to selectively engage countershaft gearelements with torque transmitting gears on the mainshaft.

It is known design practice to use a dual clutch arrangement forselectively connecting the engine to first and second torque inputgears, sometimes called headset gears, of a countershaft transmissionmechanism. An example of a dual clutch countershaft transmissionmechanism of this type may be seen by referring to U.S. patentapplication Ser. No. 10/983,531, filed Nov. 8, 2004, entitled “DualClutch Assembly For A Heavy-Duty Automotive Powertrain.” Thatapplication is assigned to the assignee of the present invention.

A dual clutch arrangement makes it possible for a countershafttransmission mechanism to be power-shifted from one ratio to another. Asone of the dual clutches is engaged, the other is disengaged. Theengaged clutch will establish a power delivery path through thecountershaft gear elements and through the main transmission gears asone or more of the ratio change clutches are engaged. A ratio changeclutch for a main transmission gear, or a countershaft gear element thatis not involved in a given selected power flow path, can be pre-engagedin preparation for a ratio change to an adjacent ratio. When a ratiochange is initiated, the ratio change clutch for one main gear or forone countershaft gear element is disengaged, and a power flow path isestablished by the pre-engaged ratio change clutch for a second maingear or a second countershaft gear element. The dual clutches of thedual clutch arrangement are alternately engaged and disengaged(“swapped” or “traded”) thereby providing a smooth transition from oneratio to an adjacent ratio in a seamless fashion.

If a multiple gear ratio power transmission mechanism of the typepreviously described is intended for use in a powertrain for amedium-duty or heavy-duty vehicle or truck, an increased number of gearratios is required throughout the engine speed range. Typically, anincreased number of gear ratios is achieved by combining a two-speedrange gearing arrangement at a torque output portion of the multiplegear ratio transmission. Although a two-speed range gearing arrangementis typical, a range gearing arrangement adapted for an increased numberof ratios can be used if a particular application for the vehiclerequires broader ratio coverage. A two-speed range gearing arrangementwill double the number of gear ratios available in the powertrain. If amultiple ratio transmission mechanism has dual clutches, it can bepower-shifted between ratios without torque interruption between theengine and the vehicle traction wheels.

If the powertrain includes multiple speed range gearing in a torque flowpath from the multiple ratio transmission mechanism to the vehicletraction wheels, it is necessary with known heavy-duty or medium-dutypowertrains for the torque flow path to be interrupted during atransition from one range to the other. That torque interruption maydeteriorate the shift quality because of inertia forces that arecreated. Further, the time needed to execute a ratio change in the rangegearing increases the time needed to execute a shift between the highestoverall ratio for one range to the adjacent lowest overall ratio foranother range.

An example of a multiple range gearing arrangement with an eight-speedtransmission is described in a publication entitled “AutomotiveHandbook,” Third Edition, published by Bosch in 1993, page 544. Theoverall number of gear ratios in the transmission described in theHandbook is sixteen, which is twice the number of ratios available inthe transmission gearing.

SUMMARY OF THE INVENTION

The invention comprises a multiple-speed transmission mechanism withdual clutches that selectively connect an engine in a vehicle powertrainto separate torque input elements of multiple ratio transmissiongearing. A multiple-range gearing arrangement is compounded with thegears of multiple-ratio transmission gearing so that a transition may bemade from a first range to a second range without torque interruption inthe torque flow path between the engine and the traction wheels. In thecase of a two-speed range gearing arrangement, the first speed rangewould be a so-called low range and the second speed range would be aso-called high range.

In accordance with a first embodiment of the invention, a firstselectively engageable clutch in the range gearing arrangementselectively connects two elements of the planetary gearing together toestablish one ratio in the range gearing arrangement, and selectivelyconnects an element of the range gearing arrangement to the transmissionhousing to establish a reaction point for another ratio in the rangegearing arrangement. A second clutch in the range gearing arrangement isused to selectively connect an element of the multiple ratiotransmission mechanism to one element of the planetary gearing and toconnect another element of the multiple ratio transmission mechanism toanother element of the planetary gearing as a transition is made betweenthe low range and the high range in the range gearing arrangement.

In accordance with a second embodiment of the invention, a countershaftgear assembly can be used rather than planetary gearing in the rangegearing arrangement. The invention is not limited in its scope, however,to a range gearing arrangement with a planetary gear or with acountershaft gear assembly since other multiple ratio gearing could beused, depending upon the application for which the powertrain isdesigned.

In each of the disclosed embodiments of the invention, the dual clutchesbetween the torque input elements of the multiple ratio transmission andthe engine can be selectively engaged and disengaged so that each gearratio in the overall speed ratio range can be preselected as power istransferred through the powertrain with an adjacent overall ratio. Thispreselection is achieved for all power shifts between ratios regardlessof whether the ratio change occurs in a low speed range or a high speedrange. Torque interruption is avoided in both embodiments duringtransitions between the ranges.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a partial cross-sectional view of a powertrain that includes amultiple-speed, dual clutch transmission in combination with two-speedrange planetary gearing;

FIG. 1 a is an enlarged view of the two-speed range planetary gearingillustrated in FIG. 1;

FIG. 2 is a schematic drawing of a dual clutch assembly for use with thetransmission of FIG. 1;

FIG. 3 is a partial cross-sectional view of a powertrain with amultiple-ratio, four-speed transmission in combination with a two-speedrange gearing arrangement with countershaft gear elements;

FIG. 4 is an illustration of the power flow path during operation of thetransmission of FIG. 1 in the first gear ratio in the low range,together with a preselected power flow path for the second gear ratio;

FIG. 5 is an illustration of the power flow path for the transmissionillustrated in FIG. 1 during operation in the second ratio in the lowrange, together with a preselected power flow path for the third ratio;

FIG. 6 is an illustration of a power flow path for the transmission ofFIG. 1 during operation in the third ratio in the low range, togetherwith a preselected power flow path for the fourth ratio;

FIG. 7 is an illustration of the power flow path for the transmission ofFIG. 1 during operation in the fourth ratio in the low range, togetherwith a preselected power flow path for the fifth ratio;

FIG. 8 is an illustration of the power flow path for the transmission ofFIG. 1 during operation in the fifth ratio in the high range, togetherwith a preselected power flow path for the sixth ratio;

FIG. 9 is an illustration of the power flow path for the transmission ofFIG. 1 during operation in the sixth ratio in the high range, togetherwith a preselected power flow path for the seventh ratio;

FIG. 10 is an illustration of the power flow path for the transmissionof FIG. 1 during operation in the seventh ratio in the high range,together with a preselected power flow path for the eighth ratio;

FIG. 11 is an illustration of the power flow path for the transmissionof FIG. 3 during operation in the first ratio in the low range, togetherwith a preselected power flow path for the second ratio;

FIG. 12 is an illustration of the transmission of FIG. 3 showing thepower flow path during operation in the second ratio in the low range,together with the preselected power flow path for the third ratio;

FIG. 13 is an illustration of the transmission of FIG. 3 showing thepower flow path during operation in the third ratio in the low range,together with the preselected power flow path for the fourth ratio;

FIG. 14 is an illustration of the transmission of FIG. 3 showing thepower flow path during operation in the fourth ratio in the low range,together with a preselected power flow path for the fifth ratio;

FIG. 15 is an illustration of the transmission of FIG. 3 showing thepower flow path during operation in the fifth ratio in the high range;

FIG. 16 is an illustration of the transmission of FIG. 3 showing thepower flow path during operation in the sixth ratio in the high range;and

FIG. 17 is an illustration of the transmission of FIG. 3 showing thepower flow path during operation in the seventh ratio in the high range.

PARTICULAR DESCRIPTION OF EMBODIMENTS OF THE INVENTION

A first embodiment of the invention, which is illustrated in FIG. 1,comprises a multiple-speed, dual clutch transmission, generallyindicated at 10, and two-speed range gearing, generally indicated at 12.Transmission 10 has a first torque input shaft 14 and a second torqueinput shaft 16. Input shaft 16 is a sleeve shaft that surrounds inputshaft 14. Sleeve shaft 16 is journaled in a bearing cap in atransmission housing, as partly shown at 18, by tapered roller bearing20. Shaft 14 is journaled within sleeve shaft 16 by tapered rollerbearing 22.

Shaft 14 is connected to a first transmission torque input gear 24 andsleeve shaft 16 is connected to second transmission torque input gear26.

A transmission mainshaft 28 is aligned with the torque input shafts andis journaled at 30 within a bearing opening in torque input gear 24. Theright-hand end of mainshaft 28 is journaled by bearing 32 in thetransmission housing so that the shaft 28 is end supported.

Torque input gears 24 and 26 are commonly referred to as headset gears.Gear 24 meshes with countershaft gear 34, and gear 26 meshes withcountershaft gear 36. Gear 34 may have a larger pitch diameter than gear36, but the relative pitch diameters could be different depending on thegear ratios of the mainshaft gears.

Countershaft gear 36 is keyed or splined to countershaft 38, which isend-supported by tapered roller bearing 40 at its left end and bytapered roller bearing 42 at its right end. Bearings 40 and 42 aresituated in bearing openings formed in the transmission housing.

Gear 34 is part of a countershaft sleeve 42 rotatably supported aboutthe axis of countershaft 38 by bearings 44 and 46. Countershaft gearelement 46 formed on countershaft sleeve 42 is in engagement withmainshaft gear 48, which is journalled on mainshaft 28.

Reverse drive pinions 50 and 52 are mounted on reverse drivecountershaft 54. Pinion 52 is in driving engagement with mainshaft gear56, which is journaled on mainshaft 28. Countershaft gear element 58 issplined or otherwise drivably connected to countershaft 38. It mesheswith mainshaft gear 60. Countershaft gear element 62, which is splinedor otherwise drivably connected to countershaft 38, meshes withmainshaft gear 64, which is journaled on mainshaft 28.

The right-hand end of mainshaft 28 is connected drivably to sun gear 66of the two-speed range gearing 12. A planetary carrier 68 of thetwo-speed range gearing 12 is connected to or is integral with a poweroutput shaft 70, which is connected drivably through a drive shaft tovehicle traction wheels. Ring gear 72 of the two-speed range gearing 12engages planet pinions carried by the carrier 68.

Ring gear 72 is selectively connected to the transmission housingportion shown at 74 and to the carrier 68 by a high/low range gearingclutch 76. The clutch 76 may be a synchronizer type clutch, as shown inFIG. 1 and in FIG. 1 a. It comprises a shiftable clutch sleeve 78 thatengages external clutch teeth on the housing portion 74 when it isshifted in a right-hand direction. When the sleeve 78 is shifted in aleft-hand direction, it will drivably engage external clutch teethformed on the carrier 68. The clutch sleeve 78 is slidably splined to anexternally splined clutch hub 79.

The shiftable clutch sleeve 78 is connected to ring gear 72, which canshift upon movement of the synchronizer clutch sleeve 78 relative toplanet pinions on carrier 68.

Gear 64, which is journaled on mainshaft 28, has external clutch teeththat continuously drivably engage internal spline teeth on splinedsleeve 82′. External clutch teeth on clutch sleeve 82 slidably engagethe internal spline teeth on spline sleeve 82′. When clutch sleeve 82 isshifted to the left, internal teeth on clutch sleeve 82 engage externalclutch teeth on clutch element 65, which connects gear 64 to mainshaft28 and to sun gear 66. Clutch element 65 is splined or otherwiseconnected at 67 to mainshaft 28. When clutch sleeve 82 is shifted to theright, external clutch teeth 86 on carrier 68 are engaged with theinternal clutch teeth on clutch sleeve 82, which connects gear 64 tocarrier 68. The external clutch teeth on clutch sleeve 82 continuouslyengage internal spline teeth on splined sleeve 82′.

Torque input gear 24 of the headset gearing and gear 48 are rotatablysupported on mainshaft 28 and may be selectively connected to mainshaft28 by ratio change clutch 92, which comprises a clutch hub 94 thatslidably supports an internally splined clutch sleeve 96. When thesleeve 96 is shifted in a right-hand direction, external clutch teeth 98on the clutch element connected to gear 48 are engaged, thusestablishing a driving connection between mainshaft 28 and gear 48. Whenthe sleeve 96 is shifted in a left-hand direction, external teeth onclutch element 100 connected to headset gear 24 are engaged, thusestablishing a driving connection between mainshaft 28 and power inputshaft 14.

The clutch 92 may be a conventional synchronizer clutch assembly thatincludes a synchronizer ring between the clutch hub 94 and the clutchteeth 98. A similar synchronizer clutch ring may be located betweenclutch hub 94 and clutch teeth carried by element 100. A spring loadedsynchronizer detent element 102 engages a recess in clutch ring 96 toestablish a synchronizer ring force on synchronizer ring 98 or onsynchronizer ring 100.

A synchronizer clutch assembly 104, which may be of a design that issimilar to the synchronizer clutch 92, has a synchronizer ring that canbe shifted into engagement with gear 60 or gear 56, thus establishing adriving connection between either of these gears with the mainshaft 28.

A dual clutch assembly that may be used to connect drivably the enginewith the two input shafts 14 and 16 is schematically illustrated in FIG.2. An engine crankshaft is connected through drive plate 105 to clutchhousing 107, which has a clutch hub 109 journaled on the transmissionhousing. A clutch friction disk 111 is situated between a clutch piston113 and friction surface 115 formed on the clutch housing 107. Anaxially offset clutch piston portion 117 is situated adjacent frictionsurface 119 on the clutch housing 107.

Clutch friction disk 111 includes a spring damper assembly 121 with adamper hub that is splined at 123 to power input shaft 16. Similarly,clutch friction disk 119 has a spring damper assembly 125 with a hubthat is splined at 127 to power input shaft 14.

Clutch engaging springs 129 are disposed between spring retainer 131 andpiston plate 113. Thus, the friction disk 111 normally is engaged underspring pressure to establish a connection between sleeve shaft 16 andthe engine crankshaft. The piston portion 117, which is slidablyconnected at its periphery to the clutch housing 107, is moved out ofengagement with friction disk 119, which is connected to shaft 14.

A pressure chamber 131 is defined by the piston plate 114 and the clutchhousing. When the chamber 131 is pressurized, friction disk 111 isdisengaged as the friction disk 119 for shaft 14 is engaged. The headsetgears 26 and 24 thus can be connected selectively to the enginecrankshaft by pressurizing and depressurizing the pressure chamber 131.

FIGS. 4, 5, 6 and 7 schematically illustrate driving power flow pathsand preselected power flow paths through the multiple-ratio transmissionand through the range gearing during operation, respectively, in thefirst ratio of a low-speed range, second ratio of the low-speed range,third ratio of the low-speed range, and fourth ratio of the low-speedrange. In each of these figures, the current power flow path isindicated by a heavy solid line and the preselected power flow path isindicated by a dotted line.

During operation in the first ratio, the clutch for hollow input shaft16 is engaged. Torque then is distributed from gear 26 to gear element36 on the countershaft 38. Torque is transferred then through thecountershaft 38 and through countershaft gear 62 through the mainshaftgear 64. The clutch sleeve for clutch 84 is shifted to the left so thatgear 64 becomes drivably connected to mainshaft 28. The clutch sleevefor range clutch 76 is shifted to the right, thus establishing amechanical connection between ring gear 72 and the transmission housing.Ring gear 72 thus is anchored to provide an anchor point for theplanetary range gearing. The speed of the carrier 68 and the poweroutput shaft 70 then is reduced relative to the speed of the mainshaft28.

During operation in the first speed ratio in low range, the secondtransmission ratio is preselected. This is accomplished by moving thesynchronizer clutch sleeve of the synchronizer clutch 92 in a right-handdirection, thus mechanically connecting mainshaft gear 48 to themainshaft 28. The low range gearing remains unchanged. A transition fromthe first ratio to the second ratio, which may be considered to be apower shift, then is accomplished by “trading” or “swapping” clutches asthe clutch for hollow input shaft 16 is released and the clutch forpower input shaft 14 is applied. The power delivery path (dotted line)that was preselected during operation in the first ratio becomes thedriving power delivery path during operation in the second ratio. Thiscondition is illustrated in FIG. 5 (solid heavy line) where torque forinput shaft 14 is delivered to headset gear 24, which drivescountershaft gear 34 and countershaft gear 46. Countershaft gear 46delivers torque to mainshaft 28 through engaged synchronizer clutch 92,which was preselected as previously indicated. Power then is deliveredthrough the range gearing, as previously described with respect to thefirst ratio.

During operation in the second driving ratio, the third driving ratio ispreselected. This is done by shifting the synchronizer clutch sleeve forsynchronizer clutch 104 in a right-hand direction. This locks the gear60 to the mainshaft 28. Clutch sleeve 82 of the clutch 84 is in aneutral position during operation in the second and third drivingratios. It does not engage either the clutch teeth 86 or the clutchteeth on clutch element 65.

A shift from the second transmission ratio to the third transmissionratio, as seen in FIG. 6, then is achieved by trading clutches so thatthe clutch for hollow input shaft 16 is engaged and the clutch for theshaft 14 is disengaged. Engine power thus is delivered to headset gear26, which drives countershaft gear element 36 and countershaft 38.Torque then is transferred from countershaft gear element 58 tomainshaft gear 60. Synchronizer clutch 104, which was preengaged duringoperation in the second ratio, connects drivably gear 60 to themainshaft 28. The range gearing remains unchanged.

A shift from the third transmission ratio to the fourth transmissionratio, as seen in FIG. 7, results in a power flow path that waspreselected during operation in the third ratio as the power flow pathfrom the torque input shaft 16 is interrupted. Torque thus is deliveredfrom shaft 14 to headset gear 24, which is drivably connected tocountershaft gear elements 34 and 46. Mainshaft gear 48 is drivablyconnected to countershaft gear 46. Gear 48 is connected drivably tomainshaft 28 through synchronizer clutch 92 since the sleeve 96 for thesynchronizer clutch 92 was pre-engaged during operation in the thirdratio. The range ratio remains unchanged as the ring gear 72 remainsanchored to the transmission housing.

Prior to a shift to the fifth ratio in the high range, the clutch sleeve82 of the synchronizer clutch assembly 84 is shifted in a right-handdirection in preparation for a subsequent shift to the fifth ratio. Thiscauses engagement of the internal teeth of the sleeve 82 with externalteeth 86 on the carrier 68. Internal sliding spline teeth on element 82′continuously engage and are fixed to external teeth on synchronizerclutch element 65. As this engagement occurs, gear 64 is not carryingtorque.

During operation in the fifth ratio in the high range, engine torquedelivered to input sleeve shaft 16 is transferred through thecountershaft gearing to mainshaft gear 64, through internal spline teethon element 82′, through synchronizer sleeve 82 and through externalsynchronizer teeth 86 to the carrier 68. The synchronizer clutch sleeve78 will move out of engagement with external clutch teeth on thetransmission housing and into engagement with external clutch teeth onthe synchronizer element 90 when desired to pre-select sixth ratio.Element 90 is splined, as shown, to the power output shaft 70. As theclutches for the headset gears are “traded,” a torque flow path fromheadset gear 24 is interrupted and a torque flow path from headset gear26 is established. This change occurs without torque interruption as atransition is made from low range to high range.

During operation in the fifth ratio, which is the lowest ratio in thehigh range, the sixth ratio is preselected by shifting synchronizerclutch sleeve 96 to the right, which drivably connects mainshaft gear 48to mainshaft 28, as seen in FIG. 9. Following a power shift of theclutches for the shaft 14 and the sleeve shaft 16, engine power duringoperation in the fifth ratio is delivered through headset gear 26,countershaft gear 58, mainshaft gear 60 and through synchronizer clutch104 to the power output shaft 70. The power output shaft 70 is directlyconnected to the mainshaft 28 at this time since the planetary rangegearing is locked up with the 1:1 ratio by the range clutch 76.

During operation in the sixth ratio, the seventh ratio is preselectedunder zero torque conditions by shifting synchronizer clutch 104 to theright, which directly connects gear 60 to the mainshaft 28. A powershift from the sixth ratio to the seventh ratio then is achieved byengaging and disengaging (“trading”) the clutches for the input shafts14 and 16. Power then is delivered through input shaft 16, throughheadset gear 26, through countershaft gear elements 36 and 58, throughmainshaft gear 60 and then through the synchronizer clutch 104 to themainshaft 28 and the power output shaft 70 as the planetary rangegearing remains in the 1:1 ratio.

During operation in the seventh ratio, the eighth ratio is preselectedby shifting synchronizer clutch sleeve 96 for the synchronizer clutch 92in a left-hand direction to pre-condition the transmission for a directconnection between the shaft 14 and the mainshaft 28.

During reverse drive, the clutch for input shaft 14 is engaged so thatheadset gear 24 will drive countershaft gear element 34 and countershaftgear element 46. Reverse 52 is drivably connected to mainshaft gear 56(the pinions 50 and 52 are shown out of position).

A second embodiment of the invention, which is shown in FIG. 3, like thefirst embodiment of the invention shown in FIGS. 1, 1 a and 3-10, is afour-speed ratio transmission, but a countershaft range gearingarrangement is used rather than a planetary range gearing arrangement.As in the case of the first embodiment, the transmission of the secondembodiment has a pair of separate headset gears that are selectivelyconnected through dual clutches to an engine, not shown. As in the caseof the first embodiment, the second embodiment comprises a multiplespeed transmission and a separate range gearing arrangement, wherein therange gearing arrangement is used to double the number of ratiosavailable in the multiple speed transmission. The number of ratiosavailable in the overall ratio range thus is eight ratios.

As in the case of the figures showing the first embodiment, a currentpower flow path for each ratio is illustrated in FIGS. 11-17 by a heavysolid line and a preselected power flow path is illustrated by a dottedline.

In FIG. 3, the four-speed dual clutch transmission is designated byreference numeral 100 and the countershaft range gearing is designatedby numeral 102.

A first headset gear 104 is connected drivably to a first power inputshaft 106. A second headset gear 108 is connected to second power inputshaft 110, which is a sleeve shaft surrounding power input shaft 106.Bearings 112 and 114 journal the headset gears in a bearing cap 116 inthe transmission housing, not shown.

A countershaft gear element 118, which is fixed to countershaft 120,engages headset gear 108. A countershaft gear element 122, which isjournaled to countershaft 120, drivably engages headset gear 104. Gearelement 122 is part of a countershaft sleeve assembly that includescountershaft gear element 124, which is in driving engagement withmainshaft gear 126. Gear 126 is journalled on mainshaft 128.

Reverse drive gear element 130 is secured to countershaft 120. Itengages reverse drive pinion 132 mounted on a fixed reverse pinioncountershaft. Reverse drive pinion 132 engages drivably reverse gear 134journaled on mainshaft 128.

Countershaft gear element 136, which is fixed to countershaft 120,engages drivably gear 138, which is journaled on mainshaft 128.

The countershaft gear 140 is fixed to countershaft 120. It drivablyengages overdrive gear 142 journaled on mainshaft 128. A synchronizerclutch 144, which corresponds in function to the synchronizer clutch 84of FIG. 1 and FIG. 1 a, is situated between overdrive gear 142 and gear146 of the countershaft range gearing 102. Gear 146 is journaled onmainshaft 128. It can be connected drivably to the mainshaft 128 whensynchronizer clutch sleeve 148 is shifted in a right-hand direction sothat internal clutch teeth on the clutch sleeve 148 engage externalclutch teeth on the gear 146.

Clutch element 150 of the clutch 144 is splined or otherwise secured tomainshaft 128, as shown at 154. When the sleeve 148 is shifted in aleft-hand direction, element 150 is connected directly to gear 142 asthe sleeve 148 drivably engage external synchronizer clutch teeth on theelement 150.

A splined sleeve element 152 has internal teeth that are in continuousengagement with external clutch teeth on the gear 142. External splineteeth are formed on the sleeve 148, which are in continuous slidingengagement with the internal teeth on the element 152.

When the sleeve 148 is shifted in a left-hand direction, external teethon element 150 are engaged, thereby establishing a direct connectionbetween gear 142 and the mainshaft 128. When sleeve 148 is shifted in aright-hand direction, the driving connection between gear 142 andmainshaft 128 is disconnected and gear 142 becomes connected to externalclutch teeth on countershaft range gear 146. This establishes a directconnection between gear 142 and gear 146.

Gear 146 is in continuous meshing engagement with countershaft gearelement 156. Countershaft gear element 156 is drivably connected tocountershaft 158 on which is formed countershaft gear element 160.Countershaft range gearing output gear 162 is in driving engagement withgear element 160 so that torque delivered to gear 162 is transferred topower output shaft 164. A bearing assembly for supporting power outputshaft is shown at 166. The countershaft 158 is end supported by bearings168 and 170 in the countershaft range gearing housing, shown in part at172.

A high/low synchronizer range gearing clutch, which corresponds to theclutch 76 shown in FIG. 1, is indicated in FIG. 3 at 174. A clutch hub176 for the synchronizer range gearing clutch 174 is splined orotherwise secured to mainshaft 128. Clutch 174 includes a clutch sleeve178 with internal splines that engage external clutch teeth of gear 162when it is shifted in a right-hand direction, thus establishing a directconnection between mainshaft 128 and the output shaft 164. When thesleeve 178 is shifted in a left-hand direction, a direct mechanicalconnection is established between mainshaft 128 and gear 146 as clutchteeth on the gear 146 are drivably engaged by the internal splined teethof the sleeve 178.

As in the case of the embodiment of the invention shown in FIG. 1 andFIG. 1 a, the embodiment of FIG. 3 has four forward driving ratios inthe transmission gearing 100. Further, the countershaft range gearingprovides two ratio ranges so that the number of ratios in the overallpowertrain is double the number of ratios available in the transmission100. The ratio changes in each ratio range are achieved by powershifting the dual clutches for the two input shafts 106 and 110 so thatas one of the dual clutches is disengaged the other is engaged, and viceversa. The ratio changes in the transmission 100 thus are achievedwithout torque interruption during the ratio change event. Further, ashift from one ratio range to the other in the countershaft rangegearing 102 is achieved without torque interruption in the same fashionas the ratio changes in the range gearing are achieved in the embodimentof FIG. 1 and FIG. 1 a.

A synchronizer clutch 180 is located between the gears 104 and 126 inFIG. 3. This synchronizer clutch corresponds to synchronizer clutch 92in the embodiment of FIGS. 1 and 1 a. A synchronizer clutch 182 islocated between mainshaft gears 134 and 138 in FIG. 3. This synchronizerclutch corresponds to synchronizer clutch 104 in the embodiment of FIGS.1 and 1 a.

FIG. 11 shows a power flow path through the transmission of FIG. 3 whenthe transmission is conditioned for the first driving ratio and thesecond preselected driving ratio in the low speed range. Synchronizerclutch 144 is centered in a neutral condition at this time. Torque istransmitted through shaft 106 to the headset gear 104. This drivescountershaft gear element 122 and countershaft gear element 124.

Synchronizer clutch 180 is shifted to the right at this time, therebyconnecting gear 126 to the mainshaft 128. A clutch sleeve forsynchronizer clutch 174 is shifted to the left and engages externalclutch teeth on mainshaft gear 146. Torque then is transmitted throughgear 126 and through the mainshaft, through countershaft range clutch174 and then to mainshaft gear 146. This drives countershaft range gearelements 156 and 160, thereby transferring torque to the output shaft164 through gear 162. At this time, the second ratio in the low range ispreselected. This is done by shifting a clutch sleeve for clutch 182 tothe right, which connects gear 138 to the mainshaft. Countershaft gearelement 136 then is conditioned to drive gear 138 and the mainshaft 128.The condition of the range gearing remains unchanged, so torque isdelivered to the output shaft 164 from the mainshaft 128.

FIG. 12 shows a power flow path for second ratio in the low range, aswell as a preselected power flow path for a third ratio. Again,synchronizer clutch 144 is in neutral. Power is delivered during secondratio operation from torque input shaft 110 to headset gear 108, whichdrives the countershaft gears 118 and 136. Synchronizer clutch 182 isshifted to the right, which connects gear 138 to the mainshaft 128.Power then is transferred through the countershaft range gearing, aspreviously described with reference to the first and second ratios.

In the third ratio in the low speed range, as seen in FIG. 13, inputshaft 106 now becomes connected to the engine through the dual clutchassembly and input shaft 110 becomes disconnected. Torque delivered toshaft 106 during operation in the third ratio is transferred to headsetgear 104, which is directly connected to the mainshaft 128 throughsynchronizer clutch 180 as the sleeve for the synchronizer clutch 180 isshifted to the left. Torque then is delivered to the output shaft 164following the same torque flow path through the range gearing previouslydescribed. The fourth ratio is preselected by shifting the synchronizerclutch 144 to the right, which connects gear 142 to the mainshaft.

The power flow path for the fourth ratio in the low speed range is shownin FIG. 14. The dual clutches again are engaged and disengaged to “swap”torque input shafts. Torque input shaft 110 now is in the torque flowpath. Thus, torque is delivered from gear 108 to countershaft gearelement 118 and to countershaft 120. This drives countershaft gearelement 140, which becomes connected through gear 142 and synchronizerclutch 144 to the mainshaft. Torque is then delivered to the outputshaft through the same torque flow path in the range gearing previouslydescribed. The fifth ratio is preselected by shifting clutch 180 to theright, which connects mainshaft gear 126 to the mainshaft 128. Thesynchronizer clutch 174 is preconditioned for torque delivery in thefifth speed range by shifting the clutch sleeve 178 in the right-handdirection, thereby locking the mainshaft to the gear 162.

The ratio change to the fifth ratio in high range from the fourth ratioin low range establishes a power flow path seen in FIG. 15. This is thelowest ratio in the high range. FIG. 15 also shows a preselected sixthratio power flow path. When the clutch for input shaft 106 is engagedand the clutch for input shaft 110 is disengaged, torque is delivered toheadset gear 108 and through countershaft gear elements 122 and 124,thereby driving gear 126. Synchronizer clutch 180, which waspreselected, now connects gear 126 to the mainshaft. Synchronizer clutch174, which was preselected during operation in the fourth ratio, isshifted to the right, thereby locking gear 162 to the mainshaft sotorque is transferred from the mainshaft to the torque output shaftthrough the engaged clutch 174.

Subsequent ratio changes in the high range take place by transitioningbetween ratios in the four-speed transmission as the torque flow paththrough the range gearing remains unchanged. During operation in thesixth ratio, as seen in FIG. 16, torque is delivered from shaft 110 toheadset gear 108 and through countershaft gear elements 118 and 136 tomainshaft gear 138 as synchronizer clutch 182 is shifted to the right.At this time, the seventh ratio is preselected by shifting synchronizerclutch 180 in a left-hand direction, which locks the input shaft 106 tothe mainshaft 128.

The power flow path for operation in the seventh ratio in the high rangeis shown in FIG. 17. At this time, input shaft 106 is directly connectedto the mainshaft through synchronizer clutch 180, as previouslyexplained. The eighth ratio is preselected at this time by shiftingsynchronizer clutch 144 to the left, which locks the overdrive gear 142to the mainshaft.

As in the case of the embodiment shown in FIGS. 1 and 1 a, range shiftsare made in the case of the embodiment of FIG. 3 without an undesirableinterruption in torque delivery. Thus, a power shift can be made whenshifting in the low range and in the high range merely by engaging anddisengaging the dual clutches. A power shift from one range to theother, as well as a power shift in the transmission gearing, is madepossible.

Each of the embodiments uses a four-speed transmission in combinationwith high/low range gearing. A transmission having a different number ofratios, however, could be used depending upon a design choice. Also, itis possible to use range gearing with more than two ranges. Further,other known synchronizer clutch constructions, including wet clutch packdesigns, could be used. This would be particularly appropriate as asubstitute for synchronizer clutch 88.

Embodiments of the invention have been disclosed, but it will beapparent to a person skilled in the art that modifications may be madewithout departing from the scope of the invention. All suchmodifications and equivalents thereof are intended to be covered by thefollowing claims.

1. A multiple speed ratio power transmission mechanism for an enginepowered vehicle having multiple gear ratio gearing, a mainshaft andrange gearing with low and high ratio ranges; two separate countershaftassemblies including separate countershafts, and multiple mainshaftgears on the mainshaft, each countershaft assembly having at least onecountershaft gear element engageable with separate mainshaft gears; twopower input shafts; each power input shaft being drivably geared to aseparate one of the countershaft assemblies; dual clutches forselectively establishing and disestablishing driving connections betweenan engine and each power input shaft; and gear ratio change clutches forselectively establishing torque flow paths through the multiple gearratio gearing; one torque flow path from the engine to one countershaftassembly and through the multiple gear ratio gearing being establishedas another of the torque flow paths from the engine to the othercountershaft assembly is preselected; the dual clutches disestablishingthe one torque flow path and establishing the other torque flow path asthe dual clutches are respectively engaged and disengaged; the rangegearing including high and low range clutches for selectivelyestablishing the low and high ratio ranges; the low ratio range beingestablished and the high ratio range being preselected during torquedelivery through the multiple gear ratio gearing; the low ratio rangebeing disestablished and the high ratio range being established as thedual clutches are engaged and disengaged to condition the transmissionmechanism for high range operation, the high and low range clutchesbeing adapted to preselect high range operation during low rangeoperation when torque delivery through range gearing for high ratiorange operation is interrupted whereby a smooth transition from lowratio range operation to high ratio range operation is made during achange in torque delivery paths through the dual clutches.
 2. A multiplespeed ratio power transmission mechanism for an engine powered vehiclehaving multiple gear ratio gearing, a mainshaft and range gearing withlow and high ratio ranges; two separate countershaft assembliesincluding separate countershafts, and multiple mainshaft gears on themainshaft, each countershaft assembly having at least one countershaftgear element engageable with separate mainshaft gears; two power inputshafts; each power input shaft being drivably geared to a separate oneof the countershaft assemblies; dual clutches for selectivelyestablishing and disestablishing driving connections between an engineand each power input shaft; and gear ratio change clutches forselectively establishing torque flow paths through the multiple gearratio gearing; one torque flow path from the engine to one countershaftassembly and through the multiple gear ratio gearing being establishedas another of the torque flow paths from the engine to the othercountershaft assembly is preselected; the dual clutches disestablishingthe one torque flow path and establishing the other torque flow path asthe dual clutches are respectively engaged and disengaged; the rangegearing including high and low range clutches for selectivelyestablishing the low and high ratio ranges; the low ratio range beingestablished and the high ratio range being preselected during torquedelivery through the multiple gear ratio gearing; the low ratio rangebeing disestablished and the high ratio range being established as thedual clutches are engaged and disengaged to condition the transmissionmechanism for high range operation, the lowest gear ratio in the highratio range being established and highest gear ratio in the low ratiorange being disestablished when the torque flow paths through the dualclutches are changed.
 3. A multiple speed ratio power transmissionmechanism for an engine powered vehicle having multiple gear ratiogearing, a mainshaft and range gearing with low and high ratio ranges;multiple mainshaft gears on the mainshaft and two separate countershaftassemblies in torque flow paths from a torque input shaft assembly; themainshaft gears on the mainshaft and countershaft gear elements on eachcountershaft assembly being in driving engagement; two power inputclutches for separately establishing an engine power flow path to eachof the countershaft assemblies; gear ratio change clutches forselectively connecting the mainshaft gears to the mainshaft; rangegearing comprising a planetary gear unit having a sun gear connected tothe mainshaft, a carrier connected to a torque output shaft and a ringgear; a low range clutch for selectively braking the ring gear; and ahigh range clutch for selectively locking together two elements of theplanetary gear unit; one gear ratio change clutch being adapted toselectively connect a mainshaft gear to the carrier during high ratiorange operation and to connect the mainshaft gear to the sun gear duringlow ratio range operation.
 4. The transmission mechanism set forth inclaim 3 wherein the dual clutches are selectively engageable to effect atransition from operation in low ratio range to operation in high ratiorange without torque delivery interruption.
 5. The multiple speed ratiotransmission set forth in claim 4 wherein the one mainshaft gear isselectively connected to the mainshaft in a torque delivery path duringoperation in highest gear ratio in the low ratio range and lowest gearratio in the high ratio range as the dual clutches are engaged anddisengaged.
 6. The transmission mechanism set forth in claim 3 whereinoperation in lowest gear ratio in the high ratio range is preselectedduring operation in highest gear ratio in the low ratio range, a powershift from the low ratio range to the high ratio range being effectedwhen torque delivery paths through the dual clutches are exchanged.